The present invention relates to a continuously variable transmission apparatus utilized as an automatic transmission apparatus for a vehicle integrated with a toroidal type continuously variable transmission, for improving a characteristic when the vehicle is stationary or at an extremely low speed.
A toroidal type continuously variable transmission, as shown in FIGS. 6 to 8, has been researched to use as an automatic transmission apparatus for a vehicle, and is partially embodied. The toroidal type continuously variable transmission is referred to as a double cavity type, and input side disks 2, 2 are supported by surroundings of both end portions of an input shaft 1 via ball splines 3, 3. Therefore, the two input side disks 2, 2 are supported concentrically and synchronizingly rotatably. Further, an output gear 4 is supported at a surrounding of a middle portion of the input shaft 1 rotatably relative to the input shaft 1. Further, output side disks 5, 5 are engaged to both end portions of a cylindrical portion provided at a central portion of the output gear 4 respectively by splines. Therefore, the two output side disks 5, 5 are synchronizingly rotated along with the output gear 4.
Further, respective pluralities (normally, two through three pieces respectively) of power rollers 6, 6 are interposed between the respective input side disks 2, 2 and the respective output side disks 5, 5. The respective power rollers 6, 6 are respectively rotatably supported by inner side faces of trunnions 7, 7 via support shafts 8, 8 and a plurality of rolling bearings. The respective trunnions 7, 7 are swingable centering on pivot shafts 9, 9 provided at the respective trunnions 7, 7 concentrically with each other at both end portions in respective length directions (up and down direction of FIGS. 6, 8 and head and tail direction of FIG. 7). A motion of inclining the respective trunnions 7, 7 is carried out by displacing the respective trunnions 7, 7 in axial directions of the pivot shafts 9, 9 by hydraulic type actuators 10, 10 and inclined angles of all of the trunnions 7, 7 are synchronized with each other hydraulically and mechanically.
That is, when the inclined angles of the respective trunnions 7, 7 are changed in order to change a transmission ratio between the input shaft 1 and the output gear 4, the respective trunnions 7, 7 are displaced by the respective actuators 10, 10 respectively in reverse directions, for example, the power roller 6 on a right side of FIG. 8 is displaced to a lower side of the drawing and the power roller 6 on a left side of the drawing is displaced to an upper side of the drawing respectively by the same distance. As a result, directions of forces in tangential lines operated to contact portions between peripheral faces of the respective power rollers 6, 6 and inner side faces of the respective input and output side disks 2, 2 and the respective output side disks 5, 5 are changed (side slip is produced at the contact portion). Further, in accordance with the change in the directions of the forces, the respective trunnions 7, 7 are swung (inclined) in directions reverse to each other centering on the pivot shafts 9, 9 axially supported by support plates 11, 11. As a result, contact positions between the peripheral faces of the respective power rollers 6, 6 and the inner side faces of the respective input side and output side disks 2, 5 are changed and a rotational transmission ratio between the input shaft 1 and the output gear 4 is changed.
A state of charging and discharging a pressurized oil to and from the respective actuators 10, 10 is controlled by a single piece of control valve 12 regardless of a number of the respective actuators 10, 10 and movement of any single piece of trunnion 7 is fed back to the control valve 12. The control valve 12 includes a sleeve 14 displaced by a stepping motor 13 in an axial direction (left and right direction of FIG. 8, head and tail direction of FIG. 6) and a spool 15 fit to an inner diameter side of the sleeve 14 displaceably in the axial direction. Further, in rods 17, 17, connecting the respective trunnions 7, 7 and pistons 16, 16 of the actuators 10, 10, an end portion of the rod 17 belonging to any single piece of the trunnion 7 is fixed with a precess cam 18 and there is constituted a feedback mechanism for transmitting movement of the rod 17, that is, a synthesized value of a displacement amount in an axial direction and a displacement amount in a rotational direction to the spool 15 via the precess cam 18 and a link arm 19. Further, a synchronizing cable 20 is hung between the respective trunnions 7, 7 to thereby mechanically synchronize the inclined angles of the respective trunnions 7, 7 even in a failure in a hydraulic system.
In switching a speed changing state, a flow path in a predetermined direction of the control valve 12 is opened by displacing the sleeve 14 to a predetermined position compatible with a desired transmission ratio by the stepping motor 13. As a result, a pressurized oil is fed in the predetermined direction to the respective actuators 10, 10 and the respective actuators 10, 10 displace the trunnions 7, 7 in the predetermined direction. That is, in accordance with feeding the pressurized oil, the respective trunnions 7, 7 are swung centering on the respective pivot shafts 9, 9 while being displaced in axial directions of the respective pivot shafts 9, 9. Further, movement (in axial direction and swinging displacement) of any single piece of the trunnion 7 is transmitted the spool 15 via the precess cam 18 fixed to the end portion of the rod 17 and the link arm 19 to displace the spool 15 in the axial direction. As a result, in a state of displacing the trunnion 7 by a predetermined amount, the flow path of the control valve 12 is closed and the pressurized oil is stopped from charging and discharging to and from the respective actuators 10, 10.
The movement of the control valve 12 based on displacements of the trunnion 7 and came face 21 of the precess cam 18 at this occasion is as follows. First, when the trunnion 7 is displaced in the axial direction in accordance with opening the flow path of the control valve 12, as described above, by the side slip produced at the contact portions between the peripheral face of the power roller 6 and the inner side faces of the input side disk 2 and the output side disk 5, the trunnion 7 starts to be displaced to swing centering on the respective pivot shafts 9, 9. Further, in accordance with the displacement in the axial direction of the trunnion 7, the displacement of the cam face 21 is transmitted to the spool 15 via the link arm 19 and the spool 15 is displaced in the axial direction to change a state of switching the control valve 12. Specifically, the control valve 12 is switched in a direction of returning the trunnion 7 to a neutral position by the actuator 10.
Therefore, the trunnion 7 starts displacing in a reverse direction toward the neutral position immediately after being changed in the axial direction. However, so far as the displacement of the neutral portion is present, the trunnion 7 continues swinging centering on the respective pivot shafts 9, 9. As a result, a displacement of the precess cam 18 with regard to a circumferential direction of the cam face 21 is transmitted to the spool 15 via the link arm 19 and the spool 15 is displaced in the axial direction. Further, simultaneously with returning the trunnion 7 to the neutral position in a state in which the inclined angle of the trunnion 7 reaches a predetermined angle matching the desired transmission ratio, the control valve 12 is closed and the pressurized oil is stopped to charge and discharge to and from the actuator 10. As a result, the inclined angle of the trunnion 7 becomes an angle compatible with an amount of displacing the sleeve 14 in the axial direction by the stepping motor 13.
In operating the above-described toroidal type continuously variable transmission, the input side disk 2 on one side (left side of FIGS. 6, 7) is driven to rotate by a drive shaft 22 connected to a power source of an engine or the like via a press apparatus 23 of a loading cam type, as illustrated. As a result, the pair of input side disks 2, 2 supported by the both end portions of the input shaft 1 are rotated synchronizingly while being pressed in directions proximate to each other. Further, the rotation is transmitted to the respective output side disks 5, 5 via the respective power rollers 6, 6 and outputted from the output gear 4.
When power is transmitted from the respective input side disks 2, 2 to the respective output side disks 5, 5 in this way, in accordance with friction at rolling contact portions (traction portions) between the peripheral faces of the respective power rollers 6, 6 supported at the respective inner side faces and the inner side faces of the respective disks 2, 5, the respective trunnions 7, 7 are exerted with forces in axial directions of the pivot shafts 9, 9 provided at the respective both end portions. The force is referred to as so-to-speak 2 Ft and a magnitude thereof is proportional to a torque transmitted from the respective input side of the disks 2, 2 to the respective output side disks 5, 5 (or from output side disks 5, 5 to input side disks 2, 2). Further, such a force 2 Ft is supported by the respective actuators 10, 10. Therefore, in operating the toroidal type continuously variable transmission, a pressure difference between pairs of hydraulic chambers 24a, 24b present on both sides of the pistons 16, 16 constituting the respective actuators 10, 10 is proportional to the magnitude of the force 2 Ft.
In the case in which rotational speeds of the input shaft 1 and the output gear 4 are changed, first, when the speed is reduced between the input shaft 1 and the output gear 4, the respective trunnions 7, 7 are moved in the axial directions of the respective pivot shafts 9, 9 by the respective actuators 10, 10 to swing to positions shown in FIG. 7. Further, as shown by FIG. 7, the peripheral faces of the respective power rollers 6, 6 are made to be respectively brought into contact with portions of the respective input side disks 2, 2 on sides of centers of the inner side faces and portions of the respective outputs side disks 5, 5 on sides of outer peripheries of the inner side faces. On the contrary, in increasing the speed, the respective trunnions 7, 7 are swung in directions opposed to those of FIG. 7 and contrary to a state shown in FIG. 7, the respective trunnions 7, 7 are inclined such that the peripheral faces of the respective power rollers 6, 6 are respectively brought into contact with portions of the respective input side disks 2, 2 on sides of the outer peripheries and portions of the respective output side disks 5, 5 on sides of the centers of the inner side faces. A middle transmission ratio (speed ratio) is provided between the input shaft 1 and the output gear 4 when the inclined angles of the respective trunnions 7, 7 are set to middles.
Further, when the toroidal type continuously variable transmission constituted and operated as described above is integrated to an actual continuously variable transmission for an automobile, it has conventionally been proposed to constitute a continuously variable transmission apparatus by combining the toroidal type continuously variable transmission to a gear type differential unit of a planetary gear mechanism or the like. For example, U.S. Pat. No. 6,251,039 discloses a continuously variable transmission apparatus which is referred to so-to-speak geared neutral and can switch a rotational state of an output shaft to regular rotation and reverse rotation by interposing a stationary state while rotating an input shaft in one direction. FIG. 9 shows such a continuously variable transmission apparatus described in U.S. Pat. No. 6,251,039. The continuously variable transmission apparatus is constituted by combining a toroidal type continuously variable transmission 25 and a planetary gear type transmission 26. The toroidal type continuously variable transmission 25 in the apparatus is provided with an input shaft 1, a pair of input side disks 2, 2, an output side disk 5a, and a plurality of power rollers 6, 6. In the illustrated example, the output side disk 5a is constituted by a structure of butting outer side faces of a pair of the output side disks to integrate.
The planetary gear type transmission 26 is provided with a carrier 27 coupled to fix to the input shaft 1 and the input side disk 2 on one side (right side of FIG. 9). A first transmitting shaft 29 both end portions of which are respectively provided fixedly with planetary gear elements 28a, 28b is rotatably supported by a middle portion in a diameter direction of the carrier 27. Further, a second transmitting shaft 31 both end portions of which are fixedly provided with sun gears 30a, 30b is supported rotatably on a side opposed to the input shaft 1 by interposing the carrier 27 there between concentrically with the input shaft 1. Further, each of the planetary gear elements 28a, 28b and a sun gear 33 fixedly provided to a front end portion (right end portion of FIG. 9) of a hollow rotating shaft 32 a base end portion (left end portion of FIG. 9) is coupled with the output side disk 5a or the sun gear 30a fixedly provided to one end portion (left end portion of FIG. 9) of the second transmitting shaft 31 are respectively brought in mesh with each other. Further, the planetary gear element 28a on one side (left side of FIG. 9) is brought in mesh with a ring gear 35 rotatably provided at a surrounding of the carrier 27 via other planetary gear element 34.
Meanwhile, planetary gear elements 37a, 37b are rotatably supported by a second carrier 36 provided at a surrounding of the sun gear 30b fixedly provided to other end portion (right end portion of FIG. 9) of the second transmitting shaft 31. Further, the second carrier 36 is fixedly provided to a base end portion (left end portion in FIG. 9) of an output shaft 38 arranged concentrically with the input shaft 1 and the second transmitting shaft 31. Further, the respective planetary gear elements 37a, 37b are brought in mesh with each other, the planetary gear element 37a on one side is brought in mesh with the sun gear 30b, and the planetary gear element 37b on other side is brought in mesh with a second ring gear 39 provided rotatably at a surrounding of the second carrier 36, respectively. Further, the ring gear 35 and the second carrier 36 are made to be engageable and disengageable by a low speed clutch 40, and the second ring gear 39 and a fixed portion of a housing or the like are made to be engageable and disengageable by a high speed clutch 41.
In the case of the above-described continuously variable transmission apparatus shown in FIG. 9, in a so-to-speak low speed mode state connecting the low speed clutch 40 and disconnecting the high speed clutch 41, power of the input shaft 1 is transmitted to the output shaft 38 via the ring gear 35. Further, by changing a transmission ratio of the toroidal type continuously variable transmission 25, a transmission ratio as a total of the continuously variable transmission apparatus, that is, a transmission ratio between the input shaft 1 and the output shaft 38 is changed. In such a low speed mode state, the gear ratio of the total of the continuously variable apparatus is changed infinitively. That is, by adjusting the transmission ratio of the toroidal type continuously variable transmission 25, while bringing the input shaft 1 in a state of being rotated in one direction, a rotational state of the output shaft 38 can be converted to regular rotation and reverse rotation by interposing a stationary state.
Further, in running at an accelerated speed or a constant speed in such a low speed mode state, a torque (passing torque) passing the toroidal type continuously variable transmission 25 is applied from the input shaft 1 to the output shaft disk 5a via the carrier 27, the first transmitting shaft 29, the sun gear 33 and the hollow rotating shaft 32 and is applied from the output side disk 5a to the respective input side disks 2, 2 via the respective power rollers 6, 6. That is, the torque passing the toroidal type continuously variable transmission 25 in running at the accelerated speed or the constant speed is circulated in a direction in which the respective input side disks 2, 2 receive the torque from the respective power rollers 6, 6.
In contrast thereto, in a so-to-speak high speed mode state in which the low speed clutch 40 is disconnected and the high speed clutch 41 is connected, the power of the input shaft 1 is transmitted to the output shaft 38 via the first and the second transmitting shafts 29, 31. Further, by changing the transmission ratio of the toroidal type continuously variable transmission 25, the transmission ratio as the total of the continuously variable transmission apparatus is changed. In this case, the larger the transmission ratio of the toroidal type continuously variable transmission 25, the larger the transmission ratio of the total of the continuously variable transmission apparatus.
Further, in running at an accelerated state or a constant speed in such a high speed mode state, a torque passing the toroidal type continuously variable transmission 25 is applied in a direction in which the respective input side disks 2, 2 apply the torque to the respective power rollers 6, 6.
In the case of the continuously variable transmission apparatus having a structure as shown by, for example, FIG. 9 and capable of realizing a so-to-speak infinity transmission ratio for stopping the output shaft 3 while the input shaft 1 is brought into a state of being rotated, it is important to maintain the torque applied to the toroidal type continuously variable transmission 25 to a proper value in a state of extremely increasing the transmission ratio including the state of stopping the output shaft 38 in view of ensuring durability of the toroidal continuously variable transmission 25 and ensuring easiness in driving operation. Because as is apparent from a relationship of “rotational drive force=rotational speed×torque”, in the state in which the transmission ratio is extremely large and the output shaft 38 is stationary or rotated at an extremely low speed while rotating the input shaft 1, the torque (passing torque) passing the toroidal type continuously variable transmission 25 becomes larger than a torque applied to the input shaft 1. Therefore, in order to ensure the durability of the toroidal type continuously variable transmission 25 without making the toroidal type continuously variable transmission 25 large-sized, it is necessary to carry out a strict control in order to confine the torque within a proper value as described above. Specifically, in order to stop the output shaft 38 while making the torque inputted to the input shaft 1 as small as possible, a control including a drive source is needed.
Further, in the state in which the transmission ratio is extremely large, even when the transmission ratio of the toroidal type continuously variable transmission 25 is slightly changed, the torque applied to the output shaft 38 is changed by a large amount. Therefore, when the transmission ratio control of the toroidal type continuously variable transmission 25 is not carried out strictly, there is a possibility that a strange feeling is given to a driver or the driver is difficult to carry out driving operation. For example, the case of an automatic transmission for an automobile, in stopping the automobile, the stationary state is maintained while the driver depressing a brake. In such a case, when the transmission ratio control of the toroidal type continuously variable transmission 25 is not carried out strictly and a large torque is applied to the output shaft 38, a force required for depressing the brake pedal in stopping the automobile is increased and fatigue of the driver is increased. Conversely, when the transmission ratio control of the toroidal type continuously variable transmission 25 is not carried out strictly in starting and the torque applied to the output shaft 38 becomes excessively small, there is a possibility that smooth starting is not carried out or a vehicle is moved rearward in starting on an uphill. Therefore, in stopping or in running at an extremely low speed, other than controlling the torque transmitted from the drive source to the input shaft 1, the transmission ratio control of the toroidal type continuously variable transmission 25 needs to carry out strictly.
In consideration of such a point, JP-A-10-103461 discloses a structure of restricting the torque (passing torque) passing the toroidal type continuously variable transmission by directly controlling the pressure difference of the hydraulic type actuator portion for displacing the trunnion.
However, in the case of the structure as disclosed in JP-A-10-103461, the control is carried out only by the pressure difference and therefore, it is difficult to stop an attitude of the trunnion at an instance at which the passing torque coincides with the target value. Specifically, since an amount of displacing the trunnion for controlling the torque is increased, so-to-speak overshooting (and hunting accompanied thereby) at which the trunnion is not stopped but continuous displacing as it is liable to be produced at the instance at which the passing torque coincides with the target value and the control of the passing torque is not stabilized.
Particularly, in the case of the toroidal type continuously variable transmission 25 which is not provided with so-to-speak cast angle in which directions of the respective pivot shafts 9, 9 provided at the both end portions of the trunnion 7, 7 and directions of center axes of the input side and the output side respective disks 2, 5 are directions orthogonal to each other as in a general half toroidal type continuously variable transmission shown in FIGS. 6 through 8, the overshooting is liable to be produced. In contrast thereto, in the case of a structure having the cast angle as in a general full toroidal type continuously variable transmission, a force in a direction of converging the overshooting is operated and therefore, it seems that sufficient torque control can be carried out even in the structure as disclosed in JP-A-10-103461.
In view of such a situation, FIG. 10 shows an example of a structure of a continuously variable transmission apparatus capable of strictly carrying out the control of the torque (passing torque) passing the toroidal type continuously variable transmission, even in a continuously variable transmission apparatus integrated with a toroidal continuously variable transmission which is not provided with the cast angle as in the general half toroidal type continuously variable transmission. Although the continuously variable transmission apparatus shown in FIG. 10 is provided with a function similar to that of the continuously variable transmission apparatus which is shown in FIG. 9 and has been known conventionally, by devising a structure of a portion of a planetary gear type transmission 26a, integrating performance of the portion of the planetary gear type transmission 26a is improved.
A first and a second planetary gear 42, 43 respectives of which are of a double pinion type are supported by both side faces of a carrier 27a rotated along with an input shaft 1 and the pair of input side disks 2, 2. That is, the first and the second planetary gears 42, 43 are constituted by respective pairs of planetary gear elements 44a, 44b, 45a, 45b. Further, the respective planetary gear elements 44a, 44b, 45a, 45b are brought in mesh with each other. The planetary gear elements 44a, 45a on an inner diameter side are respectively brought in mesh with a first and a second sun gear 47, 48 provided fixedly to a front end portion (right end portion of FIG. 10) of a hollow rotating shaft 32a with a base end portion (left end portion of FIG. 10) coupled to the output side disk 5a and one end portion (left end portion of FIG. 10) of a transmitting shaft 46. Further, the planetary gear elements 44b, 45b on an outer diameter side are brought in mesh with a ring gear 49, respectively.
Meanwhile, planetary gear elements 51a, 51b are rotatably supported by a second carrier 36a provided at a surrounding of a third sun gear 50 fixedly provided to other end portion (right end portion of FIG. 10) of the transmitting shaft 46. Further, the second carrier 36a is fixedly provided to a base end portion (left end portion of FIG. 10) of an output shaft 38a arranged concentrically with the input shaft 1. Further, the respective planetary gear elements 51a, 51b are brought in mesh with each other. The planetary gear element 51a on an inner diameter side is brought in mesh with the third sun gear 50. The planetary gear element 51b on an outer diameter side is brought in mesh with a second ring gear 39a rotatably provided at a surrounding of the second carrier 36a. Further, the ring gear 49 and the second carrier 36a are made to be engageable and disengageable by a low speed clutch 40a and the second ring gear 39a and a fixed portion of a housing or the like are made to be engageable and disengageable by a high speed clutch 41a. 
In the case of the improved continuously variable transmission apparatus constituted in this way, in a state of connecting the low speed clutch 40a and disconnecting the high speed clutch 41a, power of the input shaft 1 is transmitted to the output shaft 38a via the ring gear 49. Further, by changing a transmission ratio of the toroidal type continuously variable transmission 25, a speed ratio eCVT as a total of the continuously variable transmission apparatus, that is, a speed ratio between the input shaft 1 and the output shaft 38a is changed. A relationship between a speed ratio eCVU of the toroidal type continuously variable transmission 25 and the speed ratio eCVT of the total of the continuously variable transmission apparatus is represented by Equation (1) as follows when a ratio of a teeth number m49 of the ring gear 49 and a teeth number m47 of the first sun gear 47 is designated by notation i1 (=m49/m47).eCVT=(eCVU+i1−1)/i1  (1)
Further, when the ratio i1 of the teeth numbers is, for example, 2, the relationship between the speed ratios eCVU and eCVT is changed as shown by a line segment α of FIG. 11
In contrast thereto, in a state in which the low speed clutch 40a is disconnected and the high speed clutch 41a is connected, power of the input shaft 1 is transmitted to the output shaft 38a via the first planetary gear 42, the ring gear 49, the second planetary gear 43, the transmitting shaft 46, the respective planetary gear elements 51a, 51b and the second carrier 36a. Further, by changing the speed ratio eCVU of the toroidal type continuously variable transmission 25, the speed ratio eCVT of the total of the continuously variable transmission apparatus is changed. A relationship between the speed ratio eCVU of the toroidal type continuously variable transmission 25 and the speed ratio eCVT of the total of the continuously variable transmission apparatus at this occasion is as shown by Equation (2), shown below. Further, in Equation (2), notation i1 designates the ratio (m49/m47) of the teeth number m49 of the ring gear 49 to the teeth number m47 of the first sun gear 47, notation i2 designates a ratio (m49/m48) of the teeth number m49 of the ring gear 49 to a teeth number m48 of the second sun gear 48 and notation i3 designates a ratio (m39/m50) of a teeth number m39 of the second ring gear 39a to a teeth number m50 of the third sun gear 50, respectively.eCVT={1/(1−i3)}×{1+(i2/i1) (eCVU−1)}  (2)
Further, a relationship between the two speed ratios eCVU, eCVT when the notation i1 is 2, i2 is 2.2 and i3 is 2.8 in the respective ratios is changed as shown by a line segment β in FIG. 11.
In the case of the continuously variable transmission apparatus constituted and operated as described above, as is apparent from the line segment α of FIG. 11, a state of so-to-speak infinity speed ratio in which the output shaft 38a is stopped while the input shaft 1 is brought into a rotated state can be created. However, in the state in which the output shaft 38a is stopped while the input shaft 1 is brought into the rotating state in this way or the output shaft 38a is rotated at an extremely low speed, as described above, the torque (passing torque) passing the toroidal type continuously variable transmission 25 becomes larger than the torque applied from the engine constituting the drive source to the input shaft 1. Therefore, in stopping or in running the vehicle at a very low speed, it is necessary to properly restrict the torque inputted from the drive source to the input shaft such that the passing torque does not become excessively large (or excessively small).
Further, in running at the very small speed, in the state in which the output shaft 38a is proximate to a state of being stopped, that is, in the state in which the transmission ratio of the continuously variable transmission apparatus is very large and the rotational speed of the output shaft 38a is considerably slower than the rotational speed of the input shaft 1, the torque applied to the output shaft 38a is considerably varied by a slight variation in the transmission ratio of the continuous variable transmission apparatus. Therefore, in order to ensure smooth driving operation, it is still necessary to property restrict the torque inputted from the drive source to the input shaft 1.
Further, in running at an accelerated speed or a constant speed in such a low speed mode state, similar to the above-described structure shown in FIG. 9, the passing torque is applied from the input shaft 1 to the output side disk 5a via the carrier 27a, the first planetary gear 42, the first sun gear 43 and the hollow rotating shaft 32a and applied further from the output side disk 5a to the respective input side disks 2, 2 via the respective power rollers 6, 6 (refer to FIG. 9). That is, in running at an accelerated speed or a constant speed, the passing torque is circulated in a direction in which the respective input side disks 2, 2 receive the torque from the respective power rollers 6, 6.
Therefore, in the case of a method and an apparatus of controlling a transmission ratio by the above-described structure, as shown by FIG. 12, the torque inputted from the drive source to the input shaft 1 is properly restricted. First, the rotational speed of the engine constituting the drive source is grossly controlled. That is, the rotational speed of the engine is restricted to a point a within a range of w of FIG. 12. Along therewith the transmission ratio of the toroidal type continuously variable transmission 25 which is needed to make the rotational speed of the input shaft 1 of the continuously variable transmission apparatus coincide with the controlled rotational speed of the engine is set. The setting operation is carried out base on Equation (1), mentioned above. That is, the torque transmitted from the engine to the input shaft 1 needs to strictly restrict by the above-described method in the so-to-speak low speed mode connecting the low speed clutch 40a and disconnecting the high speed clutch 41a. Therefore, in order to constitute the rotational speed of the input shaft 1 by a value in correspondence with the necessary rotational speed of the output shaft 38a, the transmission ratio of the toroidal type continuously variable transmission 25 is set by Equation (1), mentioned above.
Further, the pressure difference between the pair of hydraulic chambers 24a, 24b (refer to FIG. 8 and FIG. 14, mentioned later) constituting the hydraulic actuators 10, 10 for displacing the trunnions 7, 7 integrated to the toroidal type continuously variable transmission 25 in the axial direction of the pivot shafts 9, 9 is measured by a hydraulic pressure sensor 52 (refer to FIG. 2, mentioned later). The hydraulic pressure measuring operation is carried out in a state in which the rotational speed of the engine is controlled grossly (however, a state of maintaining the rotational speed constant) and the transmission ratio of the toroidal type continuously variable transmission 25 is set by Equation (1). Further, by the pressure difference calculated based on the measuring operation, the torque (passing torque) TCVU passing the toroidal type continuously variable transmission 25 is calculated.
That is, so far as the transmission ratio of the toroidal type continuously variable transmission 25 is constant, the pressure difference is proportional to the torque TCVU passing the toroidal type continuously variable transmission 25 and therefore, the torque TCVU can be calculated by the pressure difference. The reason is that as described above, the respective actuators 10, 10 support the force of 2 Ft having the magnitude proportional to the torque (=TCVU passing the toroidal type continuously variable transmission 25) transmitted from the input side disks 2, 2 to the output side disk 5a (or from the output side disk 5a to the input side disks 2, 2).
Meanwhile, the above-described torque TCVU can be calculated also by Equation (3) shown below.TCVU=eCVU×TIN/{eCVU+(i1−1)ηCVU}  (3)
In Equation (3), notation eCVU designates the transmission ratio of the toroidal type continuously variable transmission 25, notation TIN designates the torque inputted from the engine to the input shaft 1, notation i1 designates the teeth number ratio of the planetary gear transmission related to the first planetary gear 42 (a ratio of teeth number m49 of the ring gear 49 to a teeth number m47 of the first sun gear 47) and notation hCVU designates an efficiency of the toroidal type continuously variable transmission 25.
Hence, based on a torque TCVU1 in actually passing the toroidal type continuously variable transmission 25 calculated from the pressure difference and the passing torque TCVU2 calculated from Equation (3) and constituting the target, a deviation ΔT (=TCVU1−TCVU2) between the actually passing torque TCVU1 and the target value TCVU2 is calculated. Further, the speed ratio of the toroidal type continuously variable transmission 25 is controlled in a direction of nullifying the deviation ΔT (a deviation is made to be ΔT=0). Further, the deviation ΔT of the torque and the deviation of the pressure difference are brought into a proportional relationship and therefore, operation of controlling the transmission ratio is carried out not only by the deviation of the torque but also by the deviation of the pressure difference. That is, the change speed ratio controlled by the deviation of the torque and the change speed ratio controlled by the deviation of the pressure difference are the same technically.
For example, as shown in FIG. 12, consider a case in which the higher the rotational speed of the input shaft 1, the more rapidly the torque TIN for driving the input shaft 1 by the engine is changed in the lowering direction in the region of restricting the torque TCVU1 (measured value) actually passing the toroidal type continuously variable transmission 25 to the target value TCVU2. Such a characteristic of the engine is easily provided even at the low speed rotational region so far as the engine is an electronically controlled engine. In the case of having such an engine characteristic and in the case in which the measured value TCVU1 of the torque is provided with a deviation in the direction in which the respective input side disks 2, 2 receive torque from the power rollers 6, 6 (refer to FIGS. 7 through 9) in comparison with the target value TCVU2 of the same, the transmission ratio of the total of the continuously variable transmission apparatus is displaced to a decelerating side in order to increase the rotational speed of the engine to reduce the torque TIN for driving the input shaft 1. Therefore, the transmission ratio of the toroidal type continuously variable transmission 25 is changed to an accelerating side.
However, in a state in which the vehicle is stopped by depressing the brake pedal (rotational speed of output shaft=0), the transmission ratio of the toroidal type continuously variable transmission 25 is controlled within a range of capable of being absorbed by slip produced at inside of the toroidal type continuously variable transmission 25, that is, slip (creep) produced at contact portions (traction portions) between the inner side faces of the respective input side and output side disks 2, 5a and the peripheral faces of the respective power rollers 6, 6 (refer to FIGS. 7 through 9). Therefore, an allowable range capable of controlling the speed ratio stays in a range in which unreasonable force is not exerted to the control portion and is limited in comparison of the case of running at a low speed.
For example, when the target value TCVU2 is present at point a in FIG. 12 and the measured value TCVU1 is present at a point b of the drawing, there is brought about a state having the deviation in a direction in which the respective input side disks 2, 2 receive the torque from the power rollers 6, 6. Hence, the speed ratio eCVU of the toroidal type continuously variable transmission 25 is changed to the accelerating side and the speed ratio eCVT of the total of the continuously variable transmission apparatus (T/M) is changed to the decelerating side. In accordance therewith, the rotational speed of the engine is increased and the torque is reduced. On the contrary, when the measured value TCVU1 is present at a point c of the drawing, there is brought about the state of having the deviation in a direction in which the input side disks 2, 2 apply the torque to the power rollers 6, 6. In this case, contrary to the above-described case, the speed ratio eCVU of the toroidal type continuously variable transmission 25 is changed to the decelerating side and the speed ratio eCVT of the total of the continuously variable transmission apparatus (T/M) is changed to the accelerating side. In accordance therewith, the rotational speed of the engine is reduced and the torque is increased.
In the following, the above-described operation is repeatedly carried out until the torque TCVU1 actually passing the toroidal type continuously variable transmission 25 which is calculated from the pressure difference coincides with the target value. That is, when the torque TCVU1 passing the toroidal type continuously variable transmission 25 cannot be made to coincide with the target value TCVU2 by controlling to change the speed of the toroidal type continuously variable transmission 25 by a single time, the above-described operation is repeatedly carried out. As a result, the torque TIN for driving to rotate the input shaft 1 by the engine can be made to be proximate to the value by which the toque TCVU passing the toroidal continuously variable transmission 25 is constituted by the target value TCVU2. Further, such an operation is carried out automatically and in a short period of time by instruction from a microcomputer integrated to a controller of the continuously variable transmission apparatus.
Further, FIG. 13 shows a relationship among a ratio of the torque TCVU passing the toroidal type continuously variable transmission 25 to the torque TIN for driving to rotate the input shaft 1 by the engine (left side ordinate), the speed ratio eCVT of the total of the continuously variable transmission apparatus (abscissa), and the speed ratio eCVU of the toroidal type continuously variable transmission 25 (right side ordinate). A bold line a shows the relationship between the ratio of the passing torque TCVU to the drive torque TIN and the speed ratio eCVT of the total of the continuously variable transmission apparatus, and a broken line b shows the relationship between the two speed ratios eCVT and eCVU, respectively. The above-described structure restricts the speed ratio eCVU of the toroidal type continuously variable transmission 25 in order to restrict the torque TCVU1 actually passing the toroidal type continuously variable transmission 25 to the target value (TCVU2) represented by a point on the bold line a in a state in which the speed ratio eCVT of the total of the continuously variable transmission apparatus is restricted to the predetermined value.
In the case of the above-described structure, the control for restricting the torque TCVU1 actually passing the toroidal type continuously variable transmission 25 to the point on the bold line “a” as the target value TCVU2 is divided in two stages, that is, the rotational speed of the engine is grossly controlled to a rotational speed which seems to provide the target value TCVU2 and thereafter, the transmission ratio control of the toroidal type continuously variable transmission 25 is carried out in accordance with the rotational speed. Therefore, the torque TCVU1 actually passing the toroidal type continuously variable transmission 25 can be restricted to the target value TCVU2 without producing overshooting (and hunting accompanied thereby) as in the method of the above-described structure, or by restraining the overshooting to be low to a degree of not being problematic practically even when the overshooting is assumedly produced.
Further, as described above, a drive force (torque) is applied to the output shaft 38a (FIG. 10) in a state of depressing the brake pedal to stop the vehicle based on the slip produced at inside of the toroidal type continuously variable transmission 25. It is conceivable to set the magnitude of the torque compatible with the creep force produced in a general automatic transmission having a torque converter which has been spread conventionally. The reason is for preventing a driver accustomed to operation of the general automatic transmission from being given a strange feeling. Further, a direction of the torque is determined by an operating position of the operating lever provided at the driver's seat. When a forward moving direction position (D range) is selected by the operating lever, the output shaft 38a is applied with the torque in the forward moving direction and when a rearward moving direction position (R range) is selected thereby, the output shaft 38a is applied with the torque in the rearward moving direction.
Next, an explanation will be given of a circuit of a portion of controlling the speed ratio of the toroidal type continuously variable transmission 25 in order to make the torque TCVU1 actually passing the toroidal type continuously variable transmission 25 coincide with the target value TCVU2 as described above in reference to FIG. 14. The pressurized oil is made to be chargeable and dischargeable to and from the pair of hydraulic chambers 24a, 24b constituting the hydraulic actuator 10 for displacing the trunnion 7 in an axial direction (up and down direction of FIG. 14) of the pivot shafts 9, 9 (refer to FIG. 8) via the control valve 12. The sleeve 14 constituting the control valve 12 is made to be displaceable in an axial direction by the stepping motor 13 via a link arm 54 and a rod 53. The spool 15 constituting the control valve 12 is engaged with the trunnion 7 via the link arm 19, and the precess cam 18 and the rod 17 and is made to be displaceable in the axial direction in accordance with an axial direction displacement and a swinging displacement of the trunnion 7. The above-described constitution is basically the same as the transmission ratio control mechanism of the toroidal type continuously variable transmission which has conventionally been known.
Particularly in the case of the above-described structure, in addition to driving the sleeve 14 by the stepping motor 13, the sleeve 14 is made to be driven also by a hydraulic pressure difference cylinder 55. That is, a front end portion of the rod 53 a base end portion of which is coupled to the sleeve 14 is axially supported by a middle portion of the link arm 54 and a pin provided at an output portion of the stepping cylinder motor 13 or the pressure difference cylinder 55 is engaged with a long hole provided at each of both end portions of the link arm 54. When the pin at inside of the long hole provided at one end portion of the link arm 54 is pushed or pulled, the pin at inside of the long hole at other end portion thereof constitutes a fulcrum. By such a constitution, the sleeve 14 is made to be displaced in the axial direction not only by the stepping motor 13 but also by the pressure difference cylinder 55. In the case of the above-described structure, by displacement of the sleeve 14 by the pressure difference cylinder 55, the speed ratio eCVU is controlled in accordance with the torque TCVU passing the toroidal type continuously variable transmission 25.
Therefore, in the case of the above-described structure, hydraulic pressures different from each other can be introduced into a pair of hydraulic chambers 56a, 56b provided at the pressure difference cylinder 55 via a correcting control valve 57. The hydraulic pressures introduced into respective hydraulic chambers 56a, 56b are determined based on a pressure difference ΔP between hydraulic pressures PDOWN and PUP operated to insides of the pair of hydraulic chambers 24a, 24b constituting an actuator 10 and a pressure difference ΔPO of output pressures of a pair of electromagnetic valves 58a, 58b for controlling an opening degree of the correcting control valve 57. That is, opening and closing of the two electromagnetic valves 58a, 58b are calculated by a controlling apparatus (controller), not illustrated, and controlled based on an output signal outputted from the controller such that the pressure difference ΔPO between the two electromagnetic valves 58a, 58b becomes a target pressure difference in correspondence with the target torque TCVU2 of the toroidal type continuously variable transmission 25. Therefore, a spool 59 constituting the correcting control valve 57 is operated with a force in accordance with the pressure difference ΔP between the hydraulic pressures operated to insides of the hydraulic chambers 24a, 24b of the actuator 10 and the pressure difference ΔPO between output pressures of the electromagnetic valves 58a, 58b constituting the target pressure difference in correspondence with the target torque TCVU2 constituting a force thereagainst.
When the torque TCVU1 actually passing the toroidal type continuously variable transmission 25 and the target torque TCVU2 coincide with each other, that is, when the difference ΔT between the passing torque TCVU1 and the target torque TCVU2 is null, a force in accordance with the pressure difference ΔP between the pressures operated to the hydraulic chambers 24a, 24b of the actuator 10 and a force in accordance with the pressure difference ΔP between output pressures of the electromagnetic valves 58a, 58b are balanced. Therefore, the spool 59 constituting the correcting control valve 57 is disposed at a neutral position and also pressures operated to the hydraulic chambers 56a, 56b of the pressure difference cylinder 55 become equal to each other. Under the state, the spool 60 of the pressure difference cylinder 55 is disposed at a neutral position and the speed ratio of the toroidal type continuously variable transmission 25 remains unchanged (not corrected).
Meanwhile, when a difference is produced between the torque TCVU1 actually passing the toroidal type continuously variable transmission 25 and the target torque TCVU2, a balance between the force in accordance with the pressure difference ΔP between hydraulic pressures operated to the hydraulic chambers 24a, 24b of the actuator 10 and the force in accordance with the pressure difference ΔPO between output pressures of the electromagnetic valves 58a, 58b is not established. Further, in accordance with a magnitude and a direction of the difference ΔT between the passing torque TCVU1 and the target torque TCVU2, the spool 59 constituting the correcting control valve 57 is displaced in an axial direction and pertinent hydraulic pressures in accordance with the magnitude and the direction of ΔT are introduced into the hydraulic chambers 56a, 56b of the pressure difference cylinder 55. Further, a spool 60 of the pressure difference cylinder 55 is displaced in the axial direction and in accordance therewith, the sleeve 14 constituting the control valve 12 is displaced in the axial direction. As a result, the trunnion 7 is displaced in the axial directions of the pivot shafts 9, 9 and the speed ratio of the toroidal type continuously variable transmission 25 is changed (corrected) Further, the direction of changing the speed ratio and the amount of changing the speed ratio in this way are as explained in reference to FIGS. 12 to 13. Further, an amount of displacing the speed ratio of the toroidal type continuously variable transmission 25 in this way, that is, the corrected amount (amount of correcting speed ratio) is sufficiently smaller than a width of the speed ratio of the toroidal type continuously variable transmission 25. Therefore, the stroke of the spool 60 of the pressure difference cylinder 55 is made to be sufficiently smaller than a stroke of the output portion of the stepping motor 13. Further, the stroke of the spool 60 restrained to be small in this way correspond to a “mechanically limited range”.
In the case of embodying the continuously variable transmission apparatus according to the above-described structure, even when the vehicle is stationary, it is preferable to strictly carry out the speed ratio control of the toroidal type continuously variable transmission integrated to the continuously variable transmission apparatus. For example, when the vehicle is stationary, it is not preferable to apply a torque which is not intended by a driver suddenly to the output shaft at an instance of switching a state of selecting a parking range (P) or a neutral range (N) constituting a state in which the vehicle is not run (no-running condition) by a select lever provided at a driver's seat to a drive range (D) or a reverse range (R) constituting a state of running the vehicle. It is preferable that the torque applied to the output shaft at the instance is equal to null or a torque to a degree of being applied thereto at an instance of switching from the P range or the N range to the D range or the R range in a general automatic transmission.
Therefore, it is preferable that the torque applied to the output shaft is null or converged to a small value even when the state of running the vehicle is assumedly selected at the instance by controlling the speed ratio of the toroidal type continuously variable transmission also in the state in which the vehicle is not run (no-running condition). Such a control can theoretically be carried out by displacing the actuators 10, 10 integrated to the toroidal type continuously variable transmission 25 as shown by FIGS. 6 through 8 in the predetermined direction by the predetermined amount and displacing to swing the respective trunnions 7, 7 supporting the power rollers 6, 6 in the predetermined direction by the predetermined angle centering on the respective pivot shafts 9, 9.
However, a number of parts integrated to the toroidal type continuously variable transmission 25 is large, further, dimensional accuracies and integrating accuracies of a number of parts thereamong effect an influence on the transmission ratio of the toroidal type continuously variable transmission 25. Therefore, when a plurality of the toroidal type continuously variable transmissions 25 are considered, even when displacement amounts of the respective actuators 10, 10 integrated to the respective toroidal type continuously variable transmissions 25 are made to be constant, it is unavoidable to some degree that a difference (individual difference) is produced in the speed ratios realized by the respective toroidal type continuously variable transmissions 25. When a single one of the toroidal type continuously variable transmission 25 is used as is carried-out currently, the above-described individual difference hardly poses a problem unless the individual difference is remarkable. In contrast thereto, according to the continuously variable transmission apparatus constituting an object of the invention, when the running state is selected, in order to converge the torque applied to the output shaft to null or a small value, it is necessary to strictly control the transmission ratio of the toroidal type continuously variable transmission 25 to a considerable degree and the individual difference poses a problem.